Change speed gearing and control



March 12, 1940. E. A. THOMPSON ,1 2

CHANGE SPEED GEARING AND CONTROL Filgd larch 15, 1937 7 Sheds-Sheet 1 .4 x y w 24/ 3 IIIII W////// ///////x 2 "47/1, W-( "V 54 7/62. 'YZ OIZIYJJOZZ March 12, 1940.

E. A. THOMPSON CHANGE SPEED GEARING AND CONTROL Filed March 15, 1937 7 Sheets-Sheet 2 March 12, 1940. E. A. THOMPSON CHANGE SPEED GEARING AND CONTROL 7 Sheets-Sheet 4 Filed latch 15. 1937 v 21 fat/d, Evin 2:022

March-12, 1940. A, THOMPSON 2,193,524

' CHANGE SPEED GEARING AND 'con'rnon File d March 15, 1937 7 Sheets-Sheet 5 March 12, 1940. E. A. T HQ MPSON 2,193,524

= CHANGE SPEED GEARING AND CONTROL Filed M rch- 15,. 1937 7 Sheets-Sheet s To GUVERNOR T0 enema THROITLE PEDAL March 12, 1940. E. A. THOMPSON CHANGE SPEED GEARING AND CONTROL Filed March 15, 1937 7-Sheets-Sheet 7 EQ QM EN mum I ll !IIIIIIIIIII III Elma/Moo Patented Mar. 12, 1940.

UNITED STATES PATENT OFFICE CHANGE SPEED GEARING AND CONTROL Earl A. Thompson, Birmingham, Mich, assignor to General Motors Corporation, Detroit, Mich, a corporation of Delaware Application March 15, 1937, Serial No. 130,956

32 Claims.

In the second application noted above are shown improvements providing smoothness of shift from one ratio to another, and a degree of master control by operator-operable members,

yielding a high degree of safety in practical use.

In the third application I show further improvements which provide added safety features for operator domination over the normally automatically shifted ratio regime, useful during a wider range of operating conditions than in the preceding disclosures.

A major improvement in the present application is the operator-operated means by which I obtain not only a controlled clutch rate control, but also provide additional compensated rate control in one unit according to the ratio in which the .other unit is driving. This feature adds smoothness to the overall operation not believed available in prior art mechanisms of this character.

Further improvement in servo means so arranged as to tend toward non-failure is shown in my double pump drive, described further in detail.

An additional feature of value is the means whereby the oil draining from the direct coupling clutch plates is accomplished during engagement, and the further means by which the oil film therein is broken atdisengagement, with a minimum of drag effect which otherwise would cause excessive wear. Of outstanding novelty is the means whereby I obtain automatic shift in both forward driving units of my assembly by sequentially combining the pressure rise characteristics of the servo pump with the further characteristics of governor and operator-operable accelerator pedal position, wherein a full range of forward driving speed ratios is made available, depending upon the settings of the manual control means, other than the aforesaid pedal.

As I have demonstrated, controls for transmission units connected in'multiple series between the power and the load for obtaining a full range of transmission ratios require a regime wherein during a cyclic shift from lowest ratio to highest ratio or vice versa, it is necessary to execute a simultaneous or nearly simultaneous shift in multiple units. An example of this is in seriatim two-speed gearing designed to provide four forward speeds, as described in my S. N. 45.184 in particular. In such an assembly, first will be low gear in both units; second will be through low in one'unit with the other in direct drive; third will be by direct drive in the first unit, with low in the second. ,I-Iigh will be direct drive in both units. In my present application I demonstrate a form of shift control wherein predetermined sequence of alternate change in the two forward driving units yields a full range of ratios under given driving circumstances.

For a. shift from second to third, for example, a change in both units is required. The alternatives of opening the torque path in one unit, making the shift in the second unit during theno-drive or low torque period of the first, and completion of shift in the first unit; or else making the effort to establish a nearly simultaneous shift in both units are herewith demonstrated. The complications of sequence controls so as to avoid jerks and uneven and overdwelling shift intervals are also troublesome from the point of view of mechanical adjustment and retention of desired shift sequences.

In the present arrangement of my invention nearly simultaneous but practically sequential shift operation in both forward driving units is achieved, for advantages in a measured change speed interval, quickly established and completed. Inertia absorption means are also shown to provide speed change interval under torque for quietness and shockless transfer to a new driving condition.

As has been discussed in my preceding cases, automatically operable or self-changing gearing consist of two ordinary types; first, wherein the no-drive interval of shift is established by opening of a main clutch; and second, wherein the main clutch need not be manipulated, the elements within the transmission providing all of the essentials for the no-drive or low torque interval. My invention discloses the latter as an example, however, with improvements in simplicity over the first noted, and a resulting economy in the number of parts.

In the preferred construction of S. N. 45,184

noted preceding, the invention is shown embodied speed devices operating between a power and a load shaft. While the example showing is in the form of epicyclic gearings having alternate torque paths, one being a direct drive, others through gearing of planetary form, the friction elements used to connect direct drive and to set up drive through gearing, may equally be used to connect parallel clutches in common fourspeed constant mesh gearboxes, within the purview of my invention.

A feature of my invention is the simplified manner of automatic control having overriding master control means to set aside automatic selection and compel shift of drive to a desired driving ratio, through measured torque capacity proportioned to torque demand.

The interconnection of my new controls with the priorly described controls of associated prime movers and variable speed transmission and clutch units is likewise herewith disclosed and augmented for a compounded power control regime yielding correlated functional speed ratio changes wherein doubly-compounded speed effects are obtained.

Additional features of novelty in my disclosure as regards auxiliary power supply, alternate measured actuation of selected speed ratio compelling and actuating mechanism, reciprocal and coordinate automatic controls therefor, involving combinations of driver will, driving conditions, and master selection controls capable not only of superseding automatically selected speed ratio settings, but also involving proportional torque capacity measured according to torque demand, will be apparent upon inspection of the following specifications, claims and illustrated in the accompanying drawings, in which:

Figure 1 is an elevation section of the forwardneutral-reverse gear showing the drive to the servo pump system, the lubrication porting and the shifter mechanism for the primary shift gearing controls.

Figure 1a is a similar view to Figure 1 of the general transmission assembly structure with the casing broken away at the bottom.

Figure 2 shows a vertical cross section of the double drive to the servo pump system taken at 22 of Figure 1.

Figure 3 is a vertical cross section taken at 33 of Figure 1 showing the special relations of the gearshaft centers of Figure l; the double unit servo pump and drive, and the speed governor and drive for the automatic ratio shifting system.

Figure 4 shows an elevation cross section at 4-4 of Figure 1a, in which the geared drive actuation system of the rear unit is described.

Figures 5 and 5a, 5b and 50, describe the clutch plates 33-69 and 36-55 of the front and rear units of Figure 1a, which provide direct drive coupling. Figure 5a is a section of the externally keyed clutch plates 36 and 55, such as shown in section in Figure la.

Figure 6 is a schematic drawing of the entire control and shift actuation system as seen generally from the left side of the vehicle in which my installation is shown as an example. This view traces out all of the control functions, both manual and automatic; Figures '7 and 8 being enlarged views of the interlocking control lever system and the valving for the front and rear units respectively.

Figure 9 is of a schematic control system similar to Figure 6 embodying modifications of the controls and actuation members wherein a closer coupled integration over the degree of drive is maintained by the manual control elements, than in the system of Figure 6.

Figure 10 is a detailed sectional view of the manually operated control valve for the rear unit of Figure 9.

Figure 11 shows the differential valve of Figure 9 in detail.

Figure 12 is a view of the porting of the automatic pressure valve 200 of Figure 9, as in running condition.

Figure 13 is'a view similar to those of Figures 6 and 7 of the interlocking lever controls, but describing modifications by which the operator may enforce continuous drive in third speed within a predetermined speed range. Figure 14 describes the modifications in the controls at the driver's position for the modifications of Figure 13.

Figure 15 is a representation of a modification.

of the differential control valve of Figures 6, 9 and 11, wherein the relative pressure operated sleeve is omitted, and the movable valve is a unitary member.

In Fig. 1 the vertical longitudinal section of the transmission structure shows the general relationships of the gearing and driving elements, with the forward-neutral-reverse gear unit at the right.

The main clutch bell housing I is attached to the transmission casing 2 by bolts 3'. Web 4 separates the forward-reverse unit from the rest of the transmission, which is substantially as in S. N. 45,184 noted preceding, and shown also in section in Figure 1a.

The main clutch driven shaft 5 is supported in the casing 2 by ball bearing 6, and input gear I is fixed non-rotatably thereto. The splined shaft 8 of the forward-reverse unit pilots the forward end of shaft 2|, where thrust bearings l0 'and I0 deliver certain axial thrusts originating in the system of gearing and shafting to the casing 2 through bearing 6. Thrust bearing I9 is in the pilot space between shafts 5 and 8.

Driven shaft 8 carries drum II, the inner surface of which is internally toothed at I2 to form the input ring gear of the front" unit. Bearing N3 in web 4 supports shaft 8 in the casing 2.

The countershaft I5 is non-rotatably supported in casing 2 and web 4. The countershaft gear body 20 rotates on bearings I4 on the countershaft IS. The first gear element l6 of the countershaft gear body 20 constantly meshes with input gear I.

The second countershaft gear element I1 is constantly meshed with reverse idler l8 supported in casing 2. The reverse idler gear l8 may also be meshed with sliding gear I9 splined to and slidable on splines 9 of shaft 8. The teeth of l9 are meshable with teeth I of gear 1.

Bushing 69 pilots shaft 8 in shaft 5 and integral gear I.

Thrust bearing 41, similar to II) and I0 transfers thrusts between shaft 2| and shaft 50 carried in the casing on bearing 49. The extension 45 of shaft 50 acts as a carrier for planets 49 spindled on shafts 46, as shown in Figure 1a.

The transmission lubrication and servo pump rotors ill and I12 of the compound pump assembly to be described later, are driven, one by countershaft gearbody 29, the other by shaft 9. The compound pump operates constantly, whenever rotational power is applied to clutch driven shaft 5, or to shaft 8,.by virtue of gears H3 and I14, rotating with the shafts respectively.

Shaft 2| is splined to carrier 22 of the front unit, and is piloted at the front end in shaft 3,

' and at the rear in the output shaft (not shown).

Ported oil passages deliver servo and lubrication oil pressure to the various units as has been described. Shaft 2| is the power output member for the front unit. The rear unit is not shown in Figure l.

Fixed to carrier 22 are spindle shafts 23 for planet pinions 24. Pinions 24 constantly mesh with the inner toothed ring l2 of drum II, and with sun gear 25. The latter is integral with sleeve 23 rotating on bearings 21 on the sleeve of carrier 22. Pinions 24 rotate on bearings 23' on shafts 23.

Governor drive parts are shown in Figure 3 in relation to the gearing of Figure 1.

In the rear uni the arrangement of elements is as shown in S. N. 45,184, except as noted in the improvements in the present case. The elevation section of Figure la is a schematic view of the transmission structure according to the case noted.

In Figure 1a, the unitary assembly in section is given, with the identifying parts of Figure l in the forward-neutral reverse unit duplicated. The so-called front unit is that one immediately adjacent the forward-neutral-reverse gear, and consists of a planetary, two-speed gear, brake-clutch operated, the brake applied by springs, regulated by outflow of fluid pressure, and the clutch ap-' plied by fluid pressure.

Bolt 32 in Fig. 1 is a clamping means for the members 29 and 30 of the reaction drum of the front unit asshown in Figure la. The web of reaction sleeve 22 is riveted at 3| to hub 34 splined at 35 for plates 33. Plates 36 are apertured to permit passage of the bolts, as in'Fig. 5b.

The structure is; output shaft 2|, "keyed or splined to carrier 22 extended to support clutch plates 33, having spindles 23 for planets 24, reaction sun gear 25 being amxed by sleeve 26 to sage 19 opening to the cylinders from gland 289 Y to which pipe 218 leads by passage 281.

Shaft 2| is the input member for the so-called" rear unit, and has integral sun gears 31 and 38- meshing with planets 4344 respectively. Output shaft is integral with carrier '45 of planets 43; and annulus 5| which constantly meshes incarrier 54 for planets 44 which constantly mesh with annulus42 rotating with drum 33 on which brake 90 may bear.

Clutch drum 59 is keyed to rotate with shaft 2 and is splined for clutch plates 60. Springs 89 are release biasing means for clutch 50, the plates 55 being keyed to rotate with drum 39, ex-

tended at 56, in which portion clutch cylinders 15 mount pistons 16, piston pins 1'! bearing against transmission assembly are as follows.

cuts; or stop short of intersection at both points. I Figure 5b shows the detail of clutch plates 36 and 55, externally keyed to rotate with drums 28 and 33 respectively. These are preferably of steel, and preformed in conical shape, in order to assist release when fluid pressure is removed for disengagement. It has been found in practise that multiple disc clutches under pressure havean un pleasant drag unless suflicientenergy is stored in the clutch plates to break the compressed oil film through progressive shear force such as afforded by this construction. This method of release control is believed novel. 1

The action of engagement in a clutch such as at 33-33 is first for fluid pressure to flow in 218 and 13, Figure la, to'move'piston 12 against the conoidal disc spring action of clutch plates and releasing springs 88. As will be described later, the servo pumps of given speed-range capacity woricagainst a resistance afforded by the brake springs 9l--9|a in addition to the above noted spring actions, which determine the regulated example, may be low for initial drive when. the

clutch torque capacity requirements are low, and

higher when the requirements are higher. From 1 Y these initial points, the building up of clutch pressure to maximum engagement pressure proceeds on a predetermined scale commensurate with the available line pressure, conditioned by the resistances acting against the pumps.

In Figure-5scores.|i2, 63 in plate 33 are spirally -cut in sections as shown'in Figure5a, wherein a predetermined. ratio. of contact surface between" the-clutch plates, and the volume of oil which can be, retained in the scores, is established. In ternaliy with planets;43 is joined by drum 52 to the example, the innermost land of the scores does not intersect the internal teeth 6|, but the V, outermost land likewise does intersect the outer edge. The scores may serve as a flat reservoir of oil, when they clutch system 33-36 is engaged. Bolt 64 is'a clamp bolt for the drum parts, similar to bolt 32, and is also a key bolt for anchorin plates 36 against rotational motion with respect to the drum.

The ratio. shifting controls integral with the In the forward-reverse unit, slider gear l9 may occupy three positions; forward drive when clutched to teeth I of gear 1, neutral, and reverse drive when meshed with reverse idler gear I8. Yoke |00 is integral with member I04 mounted on rod |0| and controlled through arm I02 and shaft |03. This linkage is also shown in Figures 3 and 6. v

By way of illustration, in Figure 4 is shown anchored in casing 2, one end of brake element 30. Fin 92, locked by a nut 92a, engages the socket of anchor piece 9|. The opposite end 93 of brake 90 is pivoted at 90a to thrust rod I80.

This brake elementas shown consists of a single turn wrapped so that upon actuation when the unit is operating, a minimum of self-energising force exists. Pre-set or pre-energised springs, held off by the fluid pressure system and controls to be described, are effective to actuate the brake at controlled rates. The detail of the servo system is shown in Figure 6. The brake 80 of the front unit is identically constructed, and internally arranged as in Figure 6.

Clutch 55-80 as in Figure 1a. couples reaction drum 39 to shaft 2I to establish direct drive in the rear unit, and brake 90 prevents rotation of drum 39 and annulus gear 42 to establish geared drive. The clutch detail is given in Figure 5. Similarly, clutch 3338 couples clutch drum 28 to drum 34 for direct drive in the front unit, and brake 80 attached to casing 2, and worked by piston rod 280, establishes geared drive. (See Figs. 4 and 9.)

In summation, either front or rear unit ,may therefore be in direct or geared drive by alternate operation of the brake or clutch respectively, yielding four net forward speeds, as in the following examples:

It should be noted that shift lever I09 fastened to shaft I03 transfers operator shifter movements For example, with reduction ratios of 1.5 to 1 in the forward unit, and 2.25 to 1 in the rear unit, the overall lowest ratio available would be 3.375.

The next lowest ratio would be 2.25 to 1, obtained by keeping the rear unit in reduction, and shifting the front unit to direct. Now if we shift the rear unit to direct, and the front unit to geared drive, we obtain a net overall reduction of 1.5 to 1. Direct drive: in both units yields 1 to 1, all elements rotating together. It is entirely feasible to obtain all of those ratios superimposed on the reverse drive gear ratio, but is unnecessary for passenger car purposes.

In Figure 6, recessed in cylinders H in flange 29 of drum 28 are clutch pistons 12, guided by pins 13 in presser plate 14. Similarly, cylinders 15 in web 56 of drum 39 are fitted with pistons 16 guided by pins 11 in presser plate 18. Drilled passages 19 lead to cylinders H and passages 19' lead to cylinders 15.

Lubrication and Servo System The main supply of transmission lubricating oil for all three transmission units is kept in the sump H9. The main drive for the double pump is by means of element 20 to which pump rotor gear I13 is fixed, and shaft I18. The pump is operating at all times whenever either shaft 5 or shaft 8 are running by virtue of the drive transmitted through gears I13 and I14. The construction is shown in Figures 1, 2, and 3.

Passage 221 receives lubrication oil from pump line 220 shown in Figs. 9 and 12, delivering to ports 22I, 223, 224, 225 and 226, from whence itis fed under pressure to the various bearings and gear elements.

Figure 2 shows the detail of the drive to, the

servo pump. Gear I14 fixed to shaft 8 drives gear coupling I15 in which is socketed governor shaft 2 and the secondary rotor shaft I16 of the servo pump. For all forward rotations of shaft 8, the secondary pump assembly I produces positive pressure, but for negative rotations, its rotation subtracts from the net line pressure produced by primary pump assembly I19.

Gear I13 affixed to countershaft body 20 drives gear I11 of hollow shaft I18, which drives the rotor "I of primary pump unit I19.

As noted in Figure 3, the primary rotor I1I drives idler gear I82, and the secondary rotor I12 drives idler gear I83. For all rotations of the engine connected shaft 5, primary pump unit I19 furnishes positive pressure, even when shaft 8 revolves in reverse. See diagram, right, in Figure 9.

The suction space I10 of the pump feeds from pipe I9I of Figure 3 and delivers through both gear discharge orifices I84 and I85 to pressure space I95 from which the main feed is delivered to ports 202 and 204 of the automatic pressure control valve 200, shown in Figs. 9 and 12.

Pump suction is exerted at space I10 because of the well-known effect of gear pumps, and pressure is developed as long as either of the two rotors receives drive through the described gearing paths. Driving in reverse does affect slightly the positive net delivery of pressure by the pump, since the pump unit driven by gear I14 is of smaller capacity, and may rob the other unit.

The outward flow of oil is best seen in Figure 9 where pressure space I95 opens to passages I96 and I91. Valve cylinder body I98 may be built into the casing 2 or separately attached. Valve member 200 moves back and forth in ported passage I99, held by spring 20I in the down" position, as in Figure 9, and lifted by pressure furnished by the pump through milled leads I95 and I 91.

The valve member 200 is a ground fit in this passage I99 to form seats between the ports which will now be enumerated; port 202 at the lower position, open to lead I from the pump; port 203 connected to the servo pressure main 238; port 204 to the transmission lubrication main 220; port 205 to pump port I91; and port 209 to the exhaust outlet valve 2I1. Stop flange 201 affords a seat for spring 20I.

Valve member 200 has a longitudinal passage 208 out part way of boss 2, which in the lowermost position is clear of the bottom of port 202, for obtaining initial pressure lift against spring 20I. Assuming an increase of pressure in the pump, valve member 200 will rise as pressure builds up behind abutment 2I0.

As the upper edge of 208 passes the lower limit of port 203, valve member 200 has already exposed port 204, the fluid pressure from lead I91 now being delivered through port 203 to the servo pressure line. Further increase in pressure lifts 208 past port 204 whence both 204 and 203 are served by the valve member 200 in the new position.

Figure 12 shows the normal running position of the valve 200 with 205 exposed for balancing effect. A decrease in pump pressure below a minimum causes pressure above and below boss 2 to become less than the force of spring 2M, and less than the brake spring pressure so that valve 200 dumps the oil from the servo line into the lubrication main 220.

At extreme operating speeds, the pump may develop more than the required pressure. Therefore, valve 200 will go to the extreme up position, with spring 2" fully compressed. At this setting, abutment 2| lis opposite the upper limit of relief port 288.

Over pressure from lead [81 may escape direct v to port 288, returning to the transmission sump by relief valve 2". At high speed the relief actionoccurs whenever the pressure requirement is exceeded, but no disconcerting change in the operators control over the servo mechanism occurs.

Some of the advantages of this valve and porting combination are that it maintains uniform servo line pressure, yields a positive servo cutoil at a given minimum pump pressure, and protects both servo and lubrication systems against excess oil pressure. An additional feature is the utility of the leak pass 2 at low speeds for initial lubrication.

One unique. feature in this construction is the transfer of relieved servo pressure to the lubrication system. This prevents hunting and consequent slipping of the clutches and brakes in the transmission assembly; a useful commercial feature for avoiding excessive wear and heat.

The particular grouping of the porting of the automatic pressure valve in the present. combination provides a new characteristic extending the range of utility of controls provided by the valve action. The ground fitting of valve 288 in bore I88 is sufllciently loose, so that a continuous leakage of pressure occurs. The graduated effect of this leakage, coupled with the restriction passage 288, the related capacities of brake and clutch cylinders, and the force of brake springs, combine to yield a net response suitable for establishing changes of speed ratio varied by pump pressure.

For example, with not sufficient pressure in line 288 to sustain clutching in the front and rear units, their brake springs lock the brakes to establish low-low, or lowest forward driving Speed ratio. Provided the other valve controls are set to deliver pressure to the front and rear unit systems, an increase of pump pressure lifts valve 288 to' a point where boss 2| 4 blocks port 288, preceding the uncovering of port 285 by abutment 2. Spring 28l is preset to yield for a given set of pressure conditions such that when the valve 288 passes from initial to running condition, there is a rapid build-up of servo line ressure at the mediumlow speeds of the pump,

at which capacity is ample to operate the fluid I servo motors of both the front and rear units.

Therefore, if the further, ratio control valving is set to admit servo line pressure from 238 to either or both of the front and rear units, the speed ratio of drive will change up to second or to higher speed ratios, established by variations of pump pressure and controlled by automatic valve 288. I88 and I58 may be considered as preselected, with actual selection determined by valve 288. In this event, the engine speed at which the operator determines to drive, the ability of the engine to handle the existing load, and the resulting speed of shaft 8 are factors which control the critical pressure for speed ratio upshift. Inlet port 288 delivers fluid pressure from line 288 and line 213 to valve I58. Relief port 288 opening to sump, may be equipped with a 'selfloaded relief valve such as 2II in Figures 9 and 12, for metering the rate of clutch release and brake engagement, as is obvious from the con-.

struction. Further port 288 relieves the end of valve I58, freeing the movement from suction, to aflord a positive response of valve positioning In this case the settings of valves.

to governor and operator-opprated movements. If in starting off in "low-low" the operator has set the handlever 88I in .high, valve I88 is in the upper position, and at a pump pressure for example of 50 pounds, valve 288 has admitted pressure to line 288, which pressure is now available to'brake piston 2! and clutch pistons I8. In a given interval, pressure in 288 will rise until brake 88 will release, and clutch 88-88 will drive at whatever pressures are existing at the end point of brake release.

At the same time the governor operated valve I58 may be in either the right or left position, depending upon the coaction of pin Ill and lever II I, and if in the right or direct drive position, the building up of line pressure in 238 will also cause actuation of brake piston 28I and clutch pistons I2 of the front unit. It will be seen that valve 288 does, under these circumstances, afford automatic ratio upshlft. If the 'handlever 38I were in the "low" position, valve action of check valve 2I8 pivoted at 2l8', having metering port 221, and mounted to respond to' back pressure flow toward the pump from the servo motors, and to swing clear when the pump pressure is positive. As soon as pump pressure falls, the release rate of pressure from the servo system is controlled by metering port 221, while the orifice capacities of 284, and the valve space between 2 and boss 2 likewise becomes eflective.

The latter expedient cushions the downshift action by letting on the brake springs 81 and 81 gradually, preventing sudden, deceleration shock loads.

The pressure at which the valve releases the servo lines is preset by adjustment of screw fltting 2l5 in the housing, bearing against the upper end of spring 28l, which may be replaced by weaker or stronger springs as desired.

Conduit 228 connects to passage 22I in web 8 of housing 2. 011 may flow freely from 22I to annular channel 222 out in the extemal portion of hollow shaft 8 of gear I2. Drilled holes nal passages in shaft 2| furnish lubricant to the gears of both the front and rear units.

Drain out of such lubricant oil under pressure finds its way back to the sump 2I9, the closing of the direct drive elements of the clutch relieving the accumulated oil in the drums.

The forward-reverse .unit obtains oil under pressure from the passages shown. Further lubrication of the unit is by customary dip in the sump oil. The compartment construction of casing 2 makes it possible to seal the entire assembly with oil pan 28I, which acts as an oil reservoir, sealing means 2I8 providing a tight joint at all contact points.

The use of a common sump for the gear units and the servo actuators yields advantages for low leakage losses, rapid re-circulation, and maintenance of proper capacity requirements, and provides one-fill service oil replacement.

The resistance drop of the circuit of oil pres- 5 sure mains is arranged so that at slow engine speeds, as at idling, fresh oil is pumped through the transmission units, available the instant the engine starts up, since the main oil pump gear I13 is constantly driven from the main clutch driven shaft 5 through gears 1-I6.

To relieve the driver or servicer from learning new modes of operation, I arrange the servo mechanism so that when the car is standing still, the driver may warm up a cold engine by holding the customary main clutch pedal in disengaging position. This operation may be carried on without forcing oil unnecessarily through the servo system.

The novel nature of the double pump system is an essential feature ofv my invention. Since shafts 5 and 8 may rotate at different speeds, and in opposite relative directions, and since it is desirable to furnish oil pressure at all times whenever any rotation whatever is imparted to the variable speed gearing; and further that it is advantageous to arrange the mechanism so that no operators whim may interfere with the circulation of oil whenever the related portions of the gearing may be rotating under either engine rotation or vehicle motion, I have herein created an oil supply system fulfilling these requirements by combining the rotation of these shafts in a driving system to a double unit pump of the augmenting type, staged for one range of pressures for forward drive, and for another range of pressures for reverse drive, yet providing adequate pressures in both ranges. This form of pump, and drive system, is believed novel, and of unusual commercial utility.

When the car is in motion, whether or not jaw clutches 1'-I9 are engaged, the pump gears always supply servo and lubrication pressure. If the jaw clutches are disengaged by shift of the hand lever 30I to neutral, the pump still will be driven by the engine, until the main clutch is disengaged, and car motion will also drive pump gear I14.

It is not possible to withhold drive from the pump, at any time when either engine or vehicle are in motion.

The automatic speed ratio controls for the operation of the front unit comprise, first, IIO, moved by the governor toward the left in Figs. 6 and '1 as engine speed increases, and pivoted to equalizer bar III at I I2.

The opposite end of the equalizer bar II I is pivoted to toggle shifter rod H3 at I36, and near its center, notch I I4 is engaged by pin II5 set in lost motion lever H6. The latter lever rotates on shaft I20 as a center; carries spring stop H1 and also arm I23.

Opposing spring stop 9 is integral with lever I2I, also rotatable on shaft I20, and having eyelet I21. Intermediate spring I25 transmits forces in compression between levers I 2I and II 6, the spring stops H9 and H1 preventing the spring from leaving position.

Pin H8 in lever I3I may engage eyelet I21 of lever I2I rotatable on shaft I20, the function of the eyelet being to provide lost motion or limit it in the clockwise rotation of control shaft I20.

Adjustable stops I28 and I29 mounted in compartment I24 are used to limit the angular position-of lever I 2I.

Lever I32 mounted external to the compartment I24, rotatable with shaft I 20, is pivoted to rod 36I at I34, and responds to the accelerator pedal movement.

At a given position of governor link pivot II2, lever I32 may rotate about center I20 counterclockwise, rotating lever I3I and pin II8 causing arm I2I and spring stop I I0 to compress spring I25 applying an increasing force to cause corresponding motion of H6. Pin II5 thereupon exerts a leftward force upon equalizer ba'r III, and on link lever II3 attached to it at pivot I36.

An increase in governor speed will tend to shift pivot I36 to the right, lever III fulcruming at I I4, tending to counteract the above mentioned motion.

Casing extension I65 supports levers I38-I39 at pivot I31.

On fixed pivot I31 toggle arms I38 and I39 are mounted, the lower arm I38 being attached to pivot coupling I40. The engaging end of link lever II3 coacts with coupling I40 assisted by spring I stressed between guide lug I43 and the pin I40 on link lever II3. Thus a rightwardly exerted force in rod II3 acts on coupling I40 to cause arm I38 to swing counterclockwise about fixed pivot I31.

The upper arm I39 of the toggle is yoked to valve body I50 by loose pivot I44, and also at its upper end, carries weighted pivot I45.

Toggle spring I48 attached to arms I38 and I39 at I49 and I45 respectively, stores energy for snapping the valve right or left, as pivot I49 of arm I38 is moved past center with relationship to fixed pivot I31 and pivot I45 of arm I39.

When valve body I50 is in the right hand position as will be seen later, fluid pressure is admitted to hold 01f brake and engage the clutch 3338 of the front unit. When it is in the left hand position, the fluid pressure of that system is released and the front unit is put in low gear drive by springs 81-8111 actuating brake band 80, through rod 280, rocker 393 and thrust rod similar to I in Figure 4.

Plate I26 rotating with shaft I 52 is rocked by movement of externally mounted lever I5I attached to shaft I52. The latter lever is joined to rod 3I0 at pivot I53, and selection movement of hand lever 30I' of Figure 6 acts to shift lever I5I to operator-selected positions, through 3I0, pivots 3I I and 309, rod 308, clevis pivot 301, curved arm 306, shaft 305, and lever 30I.

The interaction of follower I23 of lever I23 integral with I I 6 is used to prevent downshift of the front unit when the handlever 30I is in low, cam I26 restraining clockwise motion of I23 of Figure '7. Resumption of high drive restores the coaction by moving cam I26 out of contact with follower I23, whereupon the latter is free and automatic shift action is resumed.

Irregular cam slot I55 cut in cam plate I26 actuates the valve I68 connected with the operation of the rear unit. Hook I56 acts as a stop engaging with toggle arm head I45 to prevent automatic shift to high ratio in the front unit, when the plate I26 is in reverse position. This is the automatic shift lockout mechanism.

Lever I60 pivoted at I58 moves manual valve I68 at yoke 298 and carries pin I59 which fits cam slot I55 of cam plate I26. The centers of I52 and I58 are taken with respect to the arm I58-I59 and the radii of slot distances from center I52 so that movement of arm I5I by rod 3I0 forces lever I60 to follow clockwise motion of lever I5I, actuating manual valve I68 of the rear unit, for shift to low or counterclockwise for direct.

Valve I68 moved by lever I60 through yoke 298 follows the movement of lever I 5|. The cam slot I66 which moves pin in of no is arranged to establish the valve I68 in the following positions:

'Handshllt Valve Pressure Ratio Reverse Down On Low Neutral Up Ofl High w Down On Low High Up 08 High Positioning lever I attached at pivot I6 I to governor rod I I0 is pivoted to the casing at I35. Servo controls These control linkages may be mounted on the left side of the transmission casing superimposed upon the valve casting, as indicated schematically.

with the governor mechanism'at the left; and at the right the toggle support I for pivot I31. The downward extension of the valve case casting terminates at right in guide lugs I43, between which lever link H3 is constrained to move.

' Pivot pin I58 projects toward the eye of the observer and is a mounting for lever I60.

The manual control valve I68 moves in bore I61; uppermost port, 260, relieves fluid pressure from the cylindrical space 26I, dumping the oil back into the external portion of the housing, from whence it drains back to the sump 2 I 9 of the transmission. The second port, 262, leads to the outlet of the casing, from where the oil may flow, through line 212 to the control cylinder 292 for the rear unit. The port263 below is the inlet from the servo pump, and it receives oil from passage 213 through the porting shown. The port, 264, delivers fluid pressure to passage 211, from where it is used to compensate for brake releasing action of springs 91 and 91a. Straps 283 cover cylinders 282 and 292, as retainers for the spring assemblies shown.

When the valve I68 is in the lowermost position, the pressure from the servo is admitted to port 264 and through 216-211 to cylinder 295, thus compensating for existing line pressure in 219, for regulating the torque capacity of clutch 55-60, as will be explained. When the valve I68 is in the upper position, it cuts off the pressure from the line 262212 and drains ports 264 and 262. This is accomplished when the hand lever SM is placed in either low or reverse" position on the indicator plate 302. During downshift from direct to low speed ratio in the rear unit, this port relationship is effective, but not when in neutral.

When the valve I68 is at the bottom of its stroke or in the direct driveposition 'fluid'pi-essure from 263 is admitted to port 262Iandis effective to overcome the force of springs 91-91a exerted on brake 90, as well as exert pressure upon clutch plates 55--60. Port 260 is effective to drain the rear unit cylinder 292 through port 262 when the valve I68 is moved back to the low-speed position.

Clutch feed lines 218 and 219 are shown in the schematic view of Figure 6.

Figures 6 and 8 show valve plunger I50 for the control of the front unit in closed position,

with line 218 and port 261 open to exhaust. When valve I50 shifts to the right-hand position, the front unit will go to dlrectdrive.

The mechanical movement which requires the valve to occupy the described two positions is shown in Figure 6, where loose pin I44 connects the extension of valve I50 external to the case I6I to toggle levers I39--I38. The two-positional action of the togglev mechanism has been described preceding.

Piston rod 290' works against rocker I93 pivoted at I94, movement of rocker I93 exerting thrust on thrust arm I90, which is pivoted to the free end of brake band 90 at 90a, and the reduced end of which swivels in notch I92. Rod 290 is fastened to abutment member 296, which is apertured to permit passage through it of pins 291 attached to piston 294, whereby sliding abutment 214, on fixed abutment rod 300 may receive the motion caused by fluid pressure in 295 and 292.

Especial attention is directed to the construction of the retainer strap shown in Figure 6. Springs 81, 81a, and 81b react against rescaling cap 283bolted to cylinder 282, which may be integral with transmission case 2 or the metal forming the compartment I24. Springs 81 and 81a exert their tension against piston 28I, which may slide freely on rod 280. Spring 81b normally exerts pressure against abutment member 288 raised to slide on stop rod 300 attached to cap 283. Abutment member 286 is rigidly attached to the end of rod 280, and may bear against abutment member 288 to the limit of motion allowed by the ends of rods 280 and 300.

Compensating spring 284 bears against the inner face of piston 28], seating firmly against fixed abutment 285 attached to rod 280 by a lock ring.

The normal condition of the mechanical system described preceding is to apply the pressure of the springs 81, 81a and 81b to piston 28I, which at its upward limit of motion presses directly against the extended collar of abutment 285 as shown in Figure 6, loading the brake member of the front unit on engagement.

Initial fluid pressure admitted through pipe 215 to cylinder 282 may build up rapidly, being assisted by the force stored in spring 284. When the'face of piston 28I strikes the upper end of abutment 286 at point 286a, the two springs 81 and 81a are overcome and the initial positive movement of the brake releasing system begins. The continuing application of fluid pressure to cylinder 282 thereafter causes abutment 286 to engage abutment 288, pressing spring 811). Continued movement may therefore take place until the shrouded end of rod 280 engages the adjacent end of fixed abutment rod 300.

Assuming that the servo pump may only deliver a finite pressure at any one time, movement of valve I50 to a position to connect the pump with port 215 likewise affects the net pressures existing .in line 218, available for loading pistons 12 and establishing drive in clutch 3836 of the front unit. It will be understood that during the first phase of pressure increase in cylinder 282, that the first stage of pressure during which spring 284 is loading piston 28I to move toward point 286a is the initial engagement stage of the clutch, during which a fairly rapid building up of volume is accompanied by a gentle rise in pressure in clutch-cylinders 1|. After the abutments 286 and 288 meet, the increase in pressure due to the increased resistance 'of spring 81b now brought into play causes a more sudden building up of pressure on the clutch plates 3336, so that a graduated and increasing clutch capacity can be created and sustained during the clutch engaging cycle when the brake release mechanism is operative, and thereafter when the relative motion between the sets of clutch plates ceases;

Port 262 connects pump line 213 and its port 263 with line 219 when valve I68 is in the up position of Figure 9. Port 268 opens to sump through self-loaded valve (not numbered) afllxed to the casing of compartment I24. Port 264 joins the lead 216 to compensator line 211 of the rear unit, corresponding to port 261, which joins 216 to compensator line 211" for the front unit.

In Figure 9 cam plate I26, similar to I26 of Figure 6, is mounted to rotate freely on shaft I28, and is moved by lever 5 attached to the shaft, through spring M6 and lever 4I1 pivoted on the cam plate I26 at 8. Stop pins (not numbered) retain proper distance relationships between levers M5 and 4", which are bossed to hold the ends of spring 6 in place.

Lever I, moved by the governor, is arranged to transmit its motion to lever 428 which, in construction, may be integral with, or an extension of I35, although in the schematic drawings, the two levers are shown separate, joined by a shaft pivoted on the housing. When the governor mechanism goes to high speed position, levers I35 and 428 rock counterclockwise until the flat end of 428 assumes a position to block clockwise movement of lever 1. Should the operator attempt to shift the rear unit to low by movement of lever 38I of Figure 6, at a time when governor speed is upward of 70 miles per hour, for example, lever 420 will abut lever H1, and the cam plate I26 will not move, although the handlever action will yield to the operators effort because of spring 6.

Lost motion lever 6' of Figure 9 is pivoted in the compartment I24 to swing in an arc to intersect the movement of lever I3I rocking with shaft I28 and lever I32 connected as in Figure 6, to the accelerator pedal 383. Fluid pressure from 212 and 219 is led through line 8 to apertured bushing 4H and through line 2 cored and drilled in lever II6'.

Piston 3 receives fluid pressure from this,

system, line 2 opening to cylinder 4 in which piston M3 is mounted, so that whenever the rear unit is in direct drive, the clearance distance between the end of lever I3I and the base of piston 4I3 becomes less, and the net distance moved by the pedal connected lever I3I' is less, before a resulting response of lever I I6 to pedal motion is had.

The passage of clutch fluid pressure from cylinder 282 of the front unit flows through 218, and 19 to cylinders H in drum 28. Likewise the flow from cylinder 292 of the rear unit passes through 212, 219, and 19 to cylinders 15 of drum 39.

Figure 4 is a transverse vertical section through the transmission assembly. Drum 39 of the rear unit is shown in position to be gripped by brake band 98, .the fixed end of which, III, is restrained from clockwise movement by adjustable bolt 92 through the extension of housing 2. The movable end 93 of band 98 is positioned by thrust rod I98, rocker I93 and the upper end of piston rod 298, which is attached to piston 29I fltting into cylindrical recess 292 in the housing as in Figure 6. Springs 91 and 91a bear against the piston 29I in a direction to cause brake band 98 to grip drum 39 being supported by base or cap 293. Subpiston 294 slides in compensating cylinder 295 under fluid pressure supplied from the differential valve 328 described following. The three conditions of operation of this structure are; first, fluid pressure may be introduced against the head of piston 29I to counteract the force of springs 91 and 81a and thereby release brake band 98 from drum 39; second, fluid pressure admitted to subpiston cylinder 295 may change the value of the line pressure at the moment when brake release occurs; and third, the fluid pressure may work simultaneously behind both pistons 29I and 294, providing a low line pressure at the instant of brake release, which acts to limit the clutch capacity since the degree of line pressure governs the magnitude of clutch loading.

The latter expedient is to eliminate shock during ratio downshift at light engine torques, and to proportionalize torque capacity to torque demand.

The differential valve mechanism is housed in the casing 2 behind the compartment I24, and

consists of valve 328 integral with stem 32I, pressed toward its seat by spring 322 and plunger 323, the spring and plunger force being opposed by fluid pressure, as will be described further.

As shown in Figure 8, the valve 328 is mounted to slide freely in bore 3I9, the extension 324 striking stop 325 at the upper limit of motion. The upper end of bore 3I9 is enlarged to accommodate ported sleeve 326, which may slide therein. Flange 321 of sleeve 326 is pressed toward seating with the upper edge of valve 328, by spring 328 whose retainer cap 329 fastened on the valve casing is of such inner diameter as to limit the upward travel of sleeve 326, which latter is a form of piston valve.

Annular recess 338 forms a lifter port for sleeve 326, whereby fluid pressure from line 3I8 may oppose the pressure of spring 328, and change the port opening between the lower edge of the sleeve 326 and the upper face of valve 328.

Transverse ports 33I in sleeve 326 coincide generally with large annular space 332 connecting to drilled passage 262, wherein fluid pressure from 262 may always be open to space 333, between stop 325, the interior of sleeve 326 and the upper face of valve 328.

Annular outlet space 334 is of smaller diameter than sleeve 326, thus providing a'limit of downward travel therefor. Space 334 is ported at 335 and 336, the passage 331 leading to counterbalancing annulus 338; and the passage 216 being connected to the compensator chamber 295 of the actuator for the rear unit.

Spaced between annuli 334 and 338 is the exhaust annulus 339, for relieving excess pressure from space 338. Cap 329 may be of threaded form with relationship to casing 2, for variable adjustment of the tension on spring 328, in order to predetermine that pressure in line 3 I8 at which sleeve 326 will rise, and thereby establish the clearance distance between sleeve 326 and valve 328 for closing off space 334, when valve 328 is moved by plunger 323.

The net tension in spring 322 determined by normal setting of lever I22 against plunger 323 and the degree of line pressure in space 334 from lead 264, establishes the gradation characteristic in the building up of compensation action by piston 294 in the control of line pressure'in 219 and clutch cylinders 15, as will be explained in the discussion following, descriptive of the clutch capacity control.

The construction of the brake bands 30 and 30 isidentical.

The front unit parts in the brake actuation assembly are identical with those of the rear unit, rocker 393 being the same as rocker I33, the re maining thrust rod and notch construction, likewise the same. Wherever possible, identifying numbers in series pertain to similar parts having identical functions; It is not deemed necessary to show a full section of the brake actuation mechanism of the front unit, because of the parallel identity of the parts.

The bracket 304 is shown attached to the steering column in Figure 6, forming a bearing for shaft 305 and a mount for sector indicator plate 302. Handlever 30I attached to shaft 305 swings in an are over positions corresponding to high," low," neutra and "reverse respectively, as marked in abbreviations on the sector plate 302 of H. L. N. R.

Button 3I5 is a spring-loaded latch, hand operated, to hold lever 30I in any of its operationthe head at 245, to engage spring retainer plate 241. Weight arms 245 are pivoted to hub 244 at 249 and terminate in cam ends 250 and weight ends 2 5I.

External large coil spring 252 rests against plate 241 and fits a recessed seat in the flange of hub 244. Internal coil spring 254 likewise rests against retainer plate 241 and presses traveler sleeve 255 to the left. This sleeve 255 is hollowed out to a bearing fit over the spindle end 255 of shaft 2" and may slide freely axially, as thrust by the end of spring 254 acting on fiange259.

At the external end of sleeve 255 collar 251 provides connecting means for the external mechanism to be moved by the governor, as indicated in Figure 6. Normally at rest, the assembly of governor parts is as shown in Figure 6. As applied speed increases, weights 25I of arms 248 swing about pivots 243 and cam ends 250 shift sleeve 255 against the tension of spring 254. When the sleeve has moved a predetermined distance, the seat 233 of the flange 259 of sleeve 255 abuts end of spring 252, and further increase of applied speed results in weights working against the combined stresses of springs 254 and 252. It will be seen that the relative travel of sleeve 255 for a given speed increment in the latter phase is less than in the prior stage, the governor working against an increased resistance. Variations of governor speed above a predetermined point can create no change in the external control mechanism, when the weights 25I stand at wider angles to the center line of the shaft 2.

The collar 251 is arranged to move arm 350 fixed to shaft 35l of Figure 6, which latter moves arm 352 pivoted to governor rod H at 353. In this way axial motion of sleeve 255 is converted to reciprocating motion of governor rod I I0, pivot H2 and rockin'. of equalizer bar III is accomplished.

handlever linkage are shown in Figure 5.

Accelerator pedal 303 is pivoted on the floorboard of the driver's compartment, and rod 355 is pivoted at 350 so as to be moved freely by the pedal 303. Return spring 351 serves to restore the pedal to minimum throttle position. Connection 353 is to the engine throttle.

The bracket shown provides a pivot for shaft m and lever m, the rod m hooking into hole 350, and lever 354 being attached to the shaft 355, and pivoted to rod 35I at 352. Depression of the accelerator pedal 303 will therefore exert a thrust on rod 355, lever 359 will swing clockwise, and consequently rod 35I will be pulled toward the left in Figure 6. This action rocks lever I32 throughfitted pivot I34, and shaft center I20, causing pin III! to engage eyelet I21, and rock lever I2I counterclockwise. Whenever spring I 25 is so stressed, the impulse of the driver's foot is exerted to swing lever H in the same direction, consequently tending to move pin I I5 and bar I I I to the left. This is the direction of movement to pull rod II3 to the left, which may through the toggle mechanism I38I39 snap the valve I50 controlling the front unit into the low position. Whenever the accelerator pedal 303 is depressed, the described mechanism then establishes a tendency for valve I50 to downshift.

The engine carburetor is not shown, as this is believed unnecessary.

In Figure 6, the engine accelerator pedal 303 rocks shafts 358 through rod 355 and lever 359, the shaft 358 being fixed to lever 359' pivoted to rod 353 to open and close the engine throttle (not shown). Lever 354 fixed to shaft 353 is pivoted to rod 35I, in turn pivoted to lever I32 of the ratio control apparatus. Spring 351 re tracts the throttle pedal connected system.

The following assembly of parts are mounted on the transmission casing 2. Shaft I 20, to which lever I32 is attached extends into the control compartment I24, where lever I3I is attached to the shaft. Levers H5 and HI may rotate on I20. Lever I3I carries pin II8 which fits loosely in eyelet I21 of lever I2I, and carries extension arm I23, with cam follower I23.

An adjustment placed at pivot 352 permits the service operator to set the relationship between required motion of the pedal 303 for a given effect on the mechanism controlled by rod 35 I, and that at 359'--353 maybe set to determine the movement of the engine throttle, so that a predetermined throttle pedal position provides a given stress of spring I25 followed by engage ment of abutments H3 and H1 of the levers I2I and H5, to suit the engine torque-speed curve and the running resistance of the vehicle. The adjustments may be set so as to require any predetermined throttle opening before downshift occurs, and the adjustable stop system I28-l29 enables the service operator to vary the relative Effect of the throttle motion upon the shift con- In Figure 13 is shown a modification of the control structure of Figures 6 and 7, wherein camplate I25" is arranged to occupy a position beyond the high for enabling the car driver predetermined operating speed ranges. Slot I55 in camplate I25" is extended to a new position such that it may rotate counterclockwise between H" and 3rd without moving pin I59 of arm I50 and therefore allow the rear unit valve I58 controlled by that arm to remain in the upper 'Ihe relationship of the accelerator pedal and to enforce a downshift-in the front unit within position of feeding fluid pressure to the rear unit.

Extension I of camplate I26" rotates so as to intersect pin II5 of lever I23, over the rounded surface of which it may exert a camming action, shifting the pin II5 to the left, causing equalizer bar III to move to the left an equivalent distance to that caused by wide open throttle through the connections of lever I3I, pin II8, eyelet I21, lever I 2I spring I25 and lever H3 to which pin I I5 is attached, as in Figures 6 and 7.

The structure enables the operator to establish a manually enforced downshift to 3rd speed, which may be maintained indefinitely for driving in gradients, with engine braking, and is also useful for acceleration demand other than provided by throttle pedal interaction control with the governor mechansm.

At speeds above 65 miles per hour, for example, the governor may move pivot II2 so far to the left that this enforced downshift action cannot occur, and the drive will then remain in direct in both units.

The modification in the manual controls required to accommodate for the extra motion of the enforced third shift is shown in Figure 14, wherein sector plate 302 is extended to a new position marked third. The movement of the hand lever 30I from "H" to the new positionin no way affects the action of the valve I60 controlling the rear unit, since slot I55 of even radius with that of the slot I55 to the point where fluid pressure admission to the rear unit is established. I

Controls similar to the above have been described and discussed in the preceding cases noted in the superscription of the present specification.

The governor connected linkage has not, however, been put out of action, but is still able to prevent abuse of engine and transmission, in

that at an extreme high speed governor position, corresponding to a predetermined car speed of say, 63 miles per hour, point II2 can move far enough to the left to shift the valve I50 back to high, or direct drive in the automatic unit.

This yields a selective effect of engine braking and acceleration available to the car driver within definite speed ranges of engine and vehicle, wherein neither engine nor transmission mechanism may be abused, and permits the driver to establish a fixed reduction ratio for gradient work where torque rather than fuel economy is desired;

It will be noted that after. an excess speed upshift to high compelled by the governor, when the setting is for enforced third drive; the control mechanism will reset to third when the governor speed falls, and drive in "third" will be resumed, requiring no especial attention from the car driver. Resumption of normal automatic shift in the automatic unit is accomplished by resetting the handlever 30I in its high" position.

The motion of the handlever 30I from "low position to other sector required a lost motion provision in the linkage of lever 308 to shaft I03, lever I02, and slider I04, so that after jaw clutches 1-I9 are meshed, the motion of 30I may be continued. Roller 3I2 after completing the stroke of slider I04 toward mesh of 1'-I9, may ride free of cam face I 05 of the slider, so that the lost motion provision is herein accounted for. This feature combines the two motions of rocking of I02 and the sliding of I04 on rod IOI.

In Figure 9, manual valve I08 is in the up position whence pump line pressure from 213 may flow to 212, and thence to brake piston 29I and situate:

clutch pistons 16 through the described portings.

\ The valve is shown in the "down position in Figure 10, port 263from the pump line 213 being ut oil, and line 212 being opened to exhaust passage 260, and regulating valve attached at 260' to the casing. Valve 260" is made of spring steel, of a predetermined rate characteristic, such that its exposed area bears a calculated relationship, for the purpose of controlling the released pressure fiow, thereby regulating the rates of clutch release, and brake application in the rear unit.

Figure N is an enlarged view of the construc-- tion of the plunger 323 operated by throttle connected lever I22, for manipulating the differential valve 320. Abutment member 214 in Figure 9 works against spring 91c under thrust from pin 291, similarly to the action of spring 910 in Figure 6.

The external'shell of plunger 323 is bored out internally to fit collar washer 321 which may slide on the adjacent endi\of the stem 32I of valve 320. Lock ring 340 prevents the washer 321 from further movement induced by tension in spring 322.

The first increment of accelerator pedal motion rocks lever -I22 counterclockwise, further compressing spring 322, opposing the force of fluid pressure acting on the upper face of valve 320, resulting in a graduating of the orifice between the lower lip of sleeve 326 and the valve 320, thereby restricting the pressure fiow from space 334 to outlets 216211 available to create pressure on counter piston 294 in the cylinder space between abutment or wall 295 in cylinder 292.

At full pedal, the end of stem 32I meets the inner end wall of plunger 323, and positive closure of the fiow from space 334 to the piston 294.

In Figure 15, the porting and passages correspond to those of Figure 9, the uppermost port 3I8' connecting the head of stem 3I4' to the pressure line 218 of the front unit. The ported passage 264 connects to line 238, receiving net pump output pressure. Ported space 335' is joined by lead 215 to compensator lines 211 and 211', and is cross-connected to space 338' by passage 331'. Spring 322' reacts between the lower face of valve 320, and the recessed portion of plunger 323', guided by stem 32I" of the valve.

The compensator valve 320 serves the purpose of regulating the rate of transfer of torque from the geared path of torque to the clutch coupled path. It governs the degree of loading pressure on the clutch that is taking the drive, at the instant of brake release, as will be apparent from this description.

In starting, the engine idles at a given speed. With the handlever 30I in neutral position, the spinning clutch disc and driven gearing are separated from the final drive. The operator may warm up the engine, the servo pump gear circulating the transmission case oil through the described passages.

When it is desired to put the vehicle in motion, the customary clutch pedal, ,depressed by the foot, separates the mating clutch plates, the operator shifts the handlever 30I to correspond to neutral to low shift of plate I26 of Figure 6. At this point attention is directed to avaluable adjunct for absorbing the inertia of the main clutch driven plate and connected parts. In my construction, the arrangement of clutch driven shaft 5, gears 1 and I5, and servo pump, and automatic pressure valve 200, provides a predeter- 7 mined back pressure when drive is removed from this system by opening of the main clutch.

Spring I causes valve 200 to block the flow of the pump, yet permitting oil to flow to thetransition from neutral to low." so that jaw clutch I will reduce quickly to zero speed.

Assuming forward drive synchronized, as soon as the driver relaxes the main clutch pedal, engine torque is delivered to shaft 3, and to the input annulus gear I2 of the front unit. The car load is assumed to be acting on carrier element 22 connected to shaft 2| which is the output shaft of the front unit.

With load on carrier 22, and engine torque on gear I2, a force is applied to the sun gear 25 tending to give it a retrograde rotaion.

Brake band of the front unit being normally stressed for locking by the springs 31, is prevented from slipping and the retrograde motion is stopped as torque reaction begins.

Gear 25 cannot further rotate, and cage 22 moves in the same direction as the engine connected gear, the planets 24 moving orbitally and rotationally, shaft 2| being driven at a ratio to engine speed, forwardly.

With springs 91 active to load band 90 of the rear unit for locking, the application of engine torque to shaft 2| and load to output shaft 50 gives rotation to shaft 50, which then applies torque to the final drive mechanism such as road wheels, tractor treads, air or ship propellers and the like; at low speed ratio, or reduction in both the front and rear units.

Overrunning torque when the throttle is relaxed, or when the load is driving the engine as on downgrades, is overcome by the greater torque reaction sustaining force of the brake springs, therefore drag is prevented.

The application of compensating fluid pressure to subpiston 294 of Figure 6, with the handlever 30I controlling valve I53 through the linkage shown, is such that ports 253 and 254 are connected whenever the handlever 30I is placed in the low" forward position. The sub-piston also serves an additional purpose.

A similar sub-piston 235' is used to compensate for the action of the front unit brake 80, as will be described further.

Overrunning torque is then prevented from skidding the brakes because of the excess capacity of springs 91 and 81. The car operator is then free from any possibility of coasting or freewheeling, and my method increases the factor of positive control under severe operating conditions, therefore increasing safety. Engine braking in low gear is desirable from the point of view of maneuver-ability in tramc, since a more accurate control over the slow speed positioning of the vehicle is subject directly to accelerator pedal 303.

The idling gears are protected from racing during the forward driving speed change interval, since a predetermined value of torque is always being delivered through brakes 80-90 or clutches 33-35; 55-40.

The driver may new drive at will with "low setting of lever IBM and still obtain the advantages of automatic operation. Governor 25I-255 through connections 350, 35I, 352 and 0 may exert an influence on the ratio control mechanism of Figure 6. Here the equalizer bar III, on increasing governor speeds is urged to move and lever linkages.

left and to swing counterclockwise about II 5 as a fulcrum, lever I35 restricting the movement. At a given governor speed, equivalent to 'engine speed, the pivot point II2 of bar III shifts left, lever link 3 shifts right, and the toggle ISO- I33 snaps from left to right bias, causing 'valve I50 to move to a, position to connect ports 255 and 251. This delivers fluid pressure from main outlet 238 of the servo pump and automatic pressure valve system of Figure 12, to the head of piston 28I in cylinder 282 of the front unit, and to passages 218 and I9 leading to pistons 12 in cylinders II of this unit.

Piston 28l overcomes the pressure of spring 34, strikes abutment 236a of rod 280, and disengages brake 80. Pistons 12 load presser plate ll and press the clutch discs 33-35 together, squeezing the clutch disengaging springs 88.

The transition from low to direct in the front unit'has been made, and the new ratio of drive is the reduction ratio of the rear unit only.

For conditions requiring acceleration, maneuvering, or unusual torque demand, the handlever may be kept in the low position indefinitely and the mechanism will select upshift only in the front unit, depending on governor speed and throttle pedal position, the latter providing means to aifect selection through the linkage 303, 355, 359. 364, and 36I of Figure 6, and through lever I32, pin H8, arm I2I and spring I25, which latter forces lever I I6 and pin II5 to the most leftward position allowed by pin II8 bearing against the notch of lever I2I.

This interaction is 50 arranged that for normal operation while in "low setting of handlever 30I, the movement of pin II5 to maximum allowed left position can force the automatic shift to low" w thin the low speed ranges of the governor only, which is a protection against unnecessary, longcontinued operation in the extreme low ratio gearing.

Furthermore, the cam plate I26 so limits counterclockwise motion of lever I23 about shaft I20, that as soon as the front unit shifts by governor action to direct drive. the increase in governor speed resulting from the opened throttle carries pivot II2 to the left so far that the permitted maximum leftward movement of H5 is not far enough for the driver, even by full depression of the accelerator pedal 303 to further enforce a downshift in the front unit.

Now if severe driving conditions or up-grades be met, the governor speeds may be reduced far enough so that the mechanism will be urged toward downshift.

This speed range of control is governor managed up to approximately 12 miles per hour, in

which range it is not obedient to operator will to shift to a higher ratio.

One may determine this speed range by varying the strengths and adjustment of governor springs 252-250, spring I25, biasing spring I and toggle spring ll8 of Figure 8 and the setting of the efiectiye lengths of the various rods, At descending governor speeds above car speeds of 8 miles per hour, this setting normally requires the front unit to re main in direct drive, under normal operating conditions, although the requirements vary for different vehicles, purposes, engine speed and power available, and the like.

Having considered the automatic operation in one ratio range, the operation in the second range as noted preceding will now be described.

The car operator in shifting the handlever "I to high position, may fulfill one of the stated objectives, which is to cause a nearly simultaneous but actually sequential shift in both the front and rear units, so as to provide a smooth transition from one intermediate ratio to an adjacent intermediate ratio.

After the arm I 23 departs from the cam face of cam plate I26, when the handlever 30I is put in high position, the lever II6 being urged counterclockwise by spring I26 is not further restrained, and for given governor and throttle settings, may cause the valve I60 to be moved to low position, approximately at the same time as the valve I68 is shifted to its high position. This is true, however, within a limited established governor range, and at or nearly full throttle pedal position of 303.

At increased governor speeds with relaxed throttle, there is no need for the mechanical advantage of the extra reduction speed interval of 3rd, and this downshift will not occur in the front unit, when the handlever 30I is moved from low to high, so that considering overall ratios, the operator may skip 3rd speed, and shift from 2nd to direct; that is, the valve I68 only will move.

The car operator after shifting to handlever position marked H enters the new regime of control. With the handlever 30I in high, the alternation of direct and reduction drive is at the control of the combined effect of governor speed and throttle position all the way up to a governor speed corresponding, for example, to 59 miles per hour, beyond which the effect of the throttle pedal 303 can no longer enforce a shift to 3rd, or to reduction in the front unit.

Full depression of the accelerator pedal 303 when driving at below 59 miles per hour causes the arm H6 to move left using pivot II2 as a momentary fulcrum, and move link lever II3 left, snapping toggle I38-I38 left, and shifting valve I50 left, closing port 266 and opening ports 268-268 connecting port 261'to the sump. This drains clutch cylinders 1I, line 218, cylinder 282 and line 214; springs 81 applying brake 60 to drum 28; eventually stopping the rotation of the drum, and sungear which serves as the reaction element for the gearing of the front unit, establishing thereupon the reduction drive.

At all governor speeds above 59 miles per hour, the pivot point H2 is moved too far left for the pin II5 to further enforce a downshift, since lever H6 is receiving the maximum eifort deliverable by spring I25; the lever system I2III6 being prevented from further clockwise rotation by stop I28.

When the speed drifts below the given 59 miles per hour point, the ability of pin iii to enforce a downshift is resumed, as will be understood from the foregoing description.

On severe up-grades which impose a heavy torque on the engine, registered as a decrease in governor speed, and with open throttle recording the operators torque demand, the downshift will occur within a definite speed range such that smooth transition from direct to reduction in the front unit occurs, and such that the available torque is in general proportional to the torque demand. On level roads, the relative points of pivots H2 and IIS as mutually acting fulcra wi therefore determine a different response characteristic by the engine's power to sustain a given speed for the existing load, conditioned by the accelerator pedal position. On downgrades, with relaxed throttle, the governor action causes the front unit to go to direct drive and remain there over all driving speed ranges above a predetermined and selected point of approximately 1'1 miles per hour. In the range of 17-59 miles an hour it is possible for the car driver to establish a downshift through acceleration demand set up by depressing the accelerator pedal 303, or torque demand.

While it is useful for the governor mechanism to prevent a downshift in the front unit above a predetermined speed of 59 miles per hour, for example, it is just as useful to apply a similar safeguard to the rear unit. The construction of Figure 9 shows arm I35 pivoted at I6I to the governor operated rod H0, and to lever 420 fixed to shaft I35. Lever 420 is arranged to intersect lever 4I1 pivoted on pin 8 of cam plate I26, so that for governor speeds in excess of a predetermined speed, for example, 40 miles per hour, a shift of handlever I to low" will not move valve I68, but compress spring 6 through lever While the handlever itself may be moved, the resistance of spring M6 is felt against the hand, and the lever 4I5 will move clockwise, stressing thatspring without moving the valve I68, or the cam plate I26.

The reaction of governor springs 262 and 254 in the system comprising sleeve 255, collar 251, arm 350, shaft I, arm 352, rod H0, and notch of lever III provides a yieldable system between weights 25I, and the control parts moved by accelerator pedal 303 and handlever 30I. Any governor force transmitted back through the control linkage to the accelerator pedal 303, must pass through spring I26, therefore the operator can only feel that force up to the limit of the inherent compression force of the spring.

The travel of lever I32 of Figure 7 with respect to lever I23 from idling engine throttle position to the point where pin II8 begins to cause I2I to rock provides a range of exclusive engine throttle control for pedal 303 which may be adjusted to the power requirements, and guarantees sufficient torque development before the automatic control interaction can be initiated. Beyond this point, the range-of pedal movement is always involved with change of ratio except for the described checks and stops.

The driver may shift the cam plate I26 from high to low position at any time, except when the governor is at high speed positions as noted above. When this is done, valve I68 of the rear unit is moved down as in Figure 6, connecting ports 263 and 264, and venting port 26I, line 212, cylinder 282, line 218, and cylinders 15 of the rear unit. When the pedal is depressed, the lever I3I is rotated until pin II8 forces arm I2I to move counterclockwise against spring I25, causing pin II5 to take a position toward the left as in Figure 7, establishing a coupled relationship between pin H5 and notch II4, limiting the leftward point where the governor forces might exert an influence for upshift. When the handlever 30I is placed in the neutral position, the gear I8 and teeth I8 slide to a non-driving point, and valve I68 moves to direct drive position for the rear unit.

Upon shifting to reverse, the main clutch pedal is depressed, applying the reactive fluid pressure load of the pump and lubrication system as previously described, to absorb the inertias of the rotating parts, and gear I8 is moved along helical splines 8 from right to left as in Figure 1.

The slider I04 through cam 3I2 on arm I02 8,198,524 traverses along rod I III fixed to the casing, and

the arm I09 moves gear I9 from right'to left into mesh with idler gear I8. In shifting from reverse to neutral, the reverse gear I9 is demeshed from I8.

For ordinary passenger car work it is desired' that drive in reverse shall be at only one-reduction speed, whereas in various draft gear, logging engines, military tanks, excavators and the like it may be useful to arrange the transmission and controls so that the same range of variable speed ratios are available in forward and reverse drive, while the present application relates particularly to passenger vehicles in the examples. Therefore I show means to prevent the automatic shift from taking place when the handlever 3M is put in reverse,' ji'ind drive can only be in reverse at low ratios in both units.

When cam plate I is in reverse position;

hook I56, of Fig. 7 swings to prevent arm I39 of Fig. 8 and pivot I45 from moving to the right, preventing valve I50 from being moved to the right to direct drive position for the front unit.

The governor and throttle linkage cannot now operate toggle I38I39 with valve I59 locked in low position. As soon as the handlever is moved from reverse," hook I56 no longer restrains arm I39.

If an automatic ratio increase is desired for reverse, finger I56 may be omitted, and at a given governor speed, the front unit will change from low to direct drive. During reverse shift, the cam plate I26 through slot I55 guides pin I59 of lever I51 to rock lever I60 and valve I68 to its limit of,

clockwise travel, lifting valve I63 so that servo pressure from line 238 through port 263 cannot flow to port 262, so that the springs 91 are active to lock brake 89 of the rear unit.

The scope of my invention is believed to be broader than the specific examples of application described herein, since many of the details may be modified in many ways within the skill of one versed in the art.

The utility of my invention is in no way circumscribed to the demonstration herewith given, i. e., an automobile power control device, but its applicability to excavators, hoists, tractors and similar machines, machine tool drive, power shafting of boats, rail cars, and aeronautical vehicles is expressly stated herewith. The scope of my invention will be apparent in the statements of the appended claims.

I claim:

1. In automatic variable speed mechanisms, in combination, an engine, a speed control for said engine, multiple variable speed gearing driven by said engine and connected to a load, multiple friction devices arranged to establish continuous driving speed ratios within said gearing including direct coupling clutches adapted for alternate actuation with certain of such devices, fluid pressure means effective to engage and hold said clutches engaged while holding said alternately actuated devices released, valving controlling said means, and a connection between said valving and said engine speed control operative to regulate the rate of transfer of torque to said clutches from said devices according to the advanced or retarded positioning of said speed control when said valvingis making said fluid pressure means efiective.

2. In variable speed controls, in combination, an engine, a speed control for said engine, multiple variable speed gearing driven by said engine including a unit having two friction ratio devices alternately actuable by fluid pressure means, a valve controlling said unit of said gearing movable into two positions, one for admitting fluid pressure to, and the other for releasing fluid pressure from said means for alternate actuation of said devices, a second valve controlling by-passed fluid from said first named valve, and a connection between said speed control and said second named valve whereby the degree of pressure released by said second valve is regulated by the position of said speed control.

3. In variable speed gearing controls, in combination, multiple driving gear units serially arranged between a power and a load shaft including a first and a second unit, friction driving elements in said second unit adapted to establish one speed ratio between said shafts, regulating means controlling the rate of engagement of said elements during the establishing of drive in said speed ratio, and superimposed control means ef:- fective upon said regulating means to vary said rate of engagement according to the selected torque multiplication of said first named unit.

4. In automatic variable speed gear controls, a governor moved element, an accelerator pedal moved element, an intermediate element partaking of the motion of both said elements and joined to a speed ratiochanging device, 'a manually moved member arranged to shift speed ratio independently of said device when placed in given positions, and an extension of said member operative to intersect the movement of said intermediate element when placed in a given position other than the aforesaid given positions.

5. In combination, a variable speed transmission embodying a plurality of clutches and brakes actuable to establish variable speed ratios in said transmission, a fluid pressure system including pistons arranged to load said clutches for engagement to establish drive at higher speed ratios, spring loaded mechanism arranged to become effective upon a reduction in pressure in said system to lock said brakes and thereby establish drive at lower speed ratios, a variable speed servo pump driven by said transmission and connected to said system, a regulating valve between said.

' 6. In control for vehicle power transmissions,

in combination, an engine adapted to drive a plurality of transmission units comprising, aprimary variable speed unit, a secondary variable speed unit driven by said first unit, connected to a load shaft, automatically operable mechanism arranged to change the speed ratios afforded by said primary unit, automatically operable mechanism arranged to change the speed ratios afforded by said secondary unit, and manually controlled connecting means between said mechanisms efiective to establish an increase of speed ratio in said secondary unit prior to an increase of speed ratio in said primary unit.

7. In power drives for motor vehicles, in combination, a primary variable speed transmission unit from said servo pump, and a valve responsive to the pressure provided by said pump ar 

